Heat pump systems are well known. Such systems generally circulate a refrigerant through a closed-loop refrigerant transport system, which system typically includes a compressor, at least two heat exchangers (one interior heat exchanger and one exterior heat exchanger), and at least two expansion devices (one expansion device for each respective heat exchanger).
Heat pump systems are generally either air-source heat pumps, geothermal water-source heat pumps, or geothermal direct exchange (“DX”) heat pumps, all of which are well understood by those skilled in the art. Air-source heat pump systems are generally less efficient than geothermal heat pump systems, but the purchase/installation costs of an air-source system is generally less expensive than the purchase/installation costs of a geothermal system. However, the operational and maintenance costs of geothermal heat pump systems (with a DX system typically being the most advantageous) are generally less than those of an air-source heat pump. Thus, purchase/installation costs versus operational and maintenance costs are usually primary considerations in purchasing a specific type of heat pump.
There are several reasons an air-source heat pump is less efficient than a DX geothermal system (a DX system will be used for comparison because it is generally the most efficient heat pump system design). One is that the air-source system requires power to operate an exterior (outdoor) fan, which is not required for a DX system. Another is that an air-source heat pump is dependent on outdoor air temperatures (which can widely fluctuate) as a heat source in the winter and as a heat sink in the summer, while a DX system is dependent on sub-surface temperatures (which remain relatively constant).
Exterior air-source heat exchangers are usually comprised of finned refrigerant transport tubing, with an electric fan to augment air flow across the finned tubing. During the heating mode of operation, the circulating refrigerant is expanded prior to entering the exterior finned tubing in the exterior heat exchanger, so that both the pressure and temperature of the refrigerant fluid will drop. This drop in temperature permits the cold refrigerant to acquire heat from the air, since heat naturally flows to cold. However, when outdoor temperatures are approximately 40-50° F., or lower, naturally occurring moisture in the air is attracted to the cold finned tubing, where the temperature of the expanded and cold refrigerant is usually below the freezing point of water. Consequently, the moisture from the air freezes, creating a layer of frost/ice. As more and more moisture continues to be attracted from the air, more and more frost/ice builds up on the exterior finned refrigerant transport/heat exchange tubing. As the frost/ice builds up, the fan-augmented airflow across the exterior finned heat exchange tubing becomes impaired and adequate operational efficiencies are lost, since not enough heat can be acquired from the exterior air to provide design level heat to the interior space.
Normally, design level heat is supplied to the interior space when the heat absorbed from the exterior air by the refrigerant (circulating within the exterior heat exchange tubing) is accentuated by the compressor, which compressor takes in the cold (but still heat laden) refrigerant vapor and, via compression, increases both the pressure and the temperature of the circulating refrigerant, supplying the now hot refrigerant gas to the interior heat exchanger (which is typically also comprised of finned refrigerant transport tubing and a fan). After most of the heat is transferred to the interior air, the now cooled refrigerant fluid is circulated back to the exterior expansion device, where the process is repeated, to continuously acquire heat from the exterior air and supply the accentuated heat to the interior air.
Specifically regarding heating mode operation in the winter, conventional air-source heat pumps may employ a periodic “defrost” cycle when outdoor temperatures are 40-50° or less. In a typical defrost cycle, a reversing valve is engaged to direct hot refrigerant gas exiting the compressor into the exterior heat exchanger, rather than into the interior heat exchanger, to provide enough heat to melt frost/ice that has built up on the exterior heat exchanger. When the reversing valve is so engaged, the heat pump system actually operates in the cooling mode, just as it normally would in the summer.
Such an instantaneous shift of the reversing valve, to place the air-source heat pump system in a defrost cycle (or actually into the cooling mode), is noisy and is hard on the compressor (contributing to accelerated wear and tear and/or to eventual compressor failure), and simultaneously effectively requires a reverse directional flow of the refrigerant fluid within both the exterior and the interior heat exchangers. Such a reverse refrigerant directional flow also modifies the pre-existing optimum refrigerant fluid charge amount in the respective interior and exterior heat exchangers, requiring additional inefficient operational time to re-establish the optimum respective charges once the defrost cycle is over and the reversing valve is switched back into a normal heating mode position.
Even worse, during a typical defrost cycle, since the air-source heat pump system is actually operating in the cooling mode, heat from the interior air is being removed and transferred into the exterior heat exchanger to help melt the frost/ice build-up. Thus, since the interior air is effectively being air-conditioned, or cooled, during a typical defrost cycle operation in the winter, supplemental (back-up) heat is normally required to be supplied to the interior air to keep it from dropping too far below a comfortable design interior temperature level. Such cooling mode operation in the winter, coupled with a requirement for supplemental heat, is both expensive for the owner of the structure and is expensive for the electric utility company supplying power, as multiple air-source heat pumps in the area tend to create “peaking” concerns for the utility. Such peaking issues require the electric utility to either provide extra generating capacity for only relatively short periods of time, or to pay other electric providers for their extra power availability, both of which are very expensive.
Various methods of removing frost/ice from outdoor refrigerant to air heat exchangers have been developed, inclusive of supplying heat to the exterior heat exchange tubing from an external heat source (which can still be expensive and troublesome), and of supplying heat to the exterior heat exchange tubing from the hot gas discharge line exiting the compressor itself. An example of the later design is found at U.S. Pat. No. 4,279,129 to Cann, et al., as assigned to the Carrier Corporation.
In the '129 patent, hot discharge gas from the compressor discharge line, during the heating mode defrost cycle of operation, is directed simultaneously both into the exterior refrigerant to air heat exchanger and into the accumulator. This is because the stated objective is to provide enough hot gas directly to the exterior heat exchanger to melt the frost/ice, which results in a condensation of the hot gas in the exterior heat exchanger so that at least some, and more than normal, liquid phase refrigerant enters the accumulator. Since there is an increased amount of liquid phase refrigerant in the accumulator, a portion of the compressor's hot gas is also bubbled into the accumulator to vaporize the liquid phase refrigerant in the accumulator.
While the subject defrost disclosure in the '129 patent will work, the design is such that: hot gas refrigerant is not disclosed as being supplied into the exterior heat exchanger, in the heating/defrost mode, through larger than normal liquid phase refrigerant transport lines, which will impair defrost abilities and will disadvantageously increase defrost period timing; there is no restriction shown in the exiting vapor transport line from the exterior heat exchanger, which will impair both defrost ability and will disadvantageously increase defrost period timing; there is no disclosure that any extra refrigerant charge is required, which will impair optimum system operation; an optimum sizing of the hot gas line by-passing the indoor heat exchanger is not disclosed, which makes it impossible to know how to optimize defrost results; and some unidentified amount of the hot gas by-passing the indoor heat exchanger is being directed into the accumulator, thereby not affording full optimum utilization of the by-passing hot gas to provide heat to interior air or to melt the frost/ice on the heat exchange tubing of the exterior heat exchanger, and thereby increasing defrost cycle operational time (which impairs both interior heat supply abilities and defrost cycle operational timing), and which unnecessarily requires hot gas to be directed and bubbled into the accumulator.
Also, since no relatively short and frequent time period for defrost cycle operation is disclosed, via an anticipated normal and customary defrost cycle operation of about four to twelve minutes with Cann et al.'s disclosure (with the defrost cycle operating only when the frost/ice build-up is relatively thick on the exterior heat exchanger's heat transfer tubing), customarily longer periods of defrost cycle operation may be required, resulting in more interior heat losses during the defrost cycle, and resulting in more condensation issues with the hot gas refrigerant used to melt the frost/ice in the outdoor heat exchange tubing (which is an additional likely reason Cann, et al. requires some hot gas to be diverted to the accumulator). Further, by diverting the hot discharge gas into both the exterior heat exchanger and the accumulator during the defrost mode of operation, there will likely be an insufficient amount of hot gas available to supply an optimum amount of heat to the interior air handler during a defrost cycle; and the subject design makes no mention of a timed operation of the fan in the exterior heat exchanger, and, of necessity as disclosed, will impose a power peaking period upon the system as a result of the defrost cycle operation as disclosed therein.
Another example of utilizing hot gas to assist in the defrost cycle of an exterior heat exchanger is found in Park's U.S. Patent Application Pub. No. US 2009/0277207 A1. In the '207 application, hot discharge gas from the compressor discharge line, during the heating mode defrost cycle of operation, is primarily directed in a quantity up to 100% solely into the exterior refrigerant to air heat exchanger. During the defrost cycle, which takes 30-100 seconds, there is no interior heating (see page 5, paragraph 0070). Park also explains that, in a case where there is an excessive quantity of frost, a three-way valve supplies the hot gas to the exterior heat exchanger only at 20-30 second intervals, during which interior heating is interrupted, so that it is difficult for the user to recognize that the heating mode has actually been fully stopped during those 20-30 second intervals (see page 5, paragraphs 0070 and 0071).
Again, while Park's design will work, Park does not provide a full unrestricted 100% hot gas flow into the exterior heat exchanger, through larger than normal liquid phase refrigerant transport lines, since the hot gas primarily goes through a conventional (small) distributor (Park's number 16) and (small) distribution tubes (Park's number 33); Park does not disclose a preferable optimum size of the hot gas by-pass line to the exterior heat exchanger; and Park provides no restriction in the vapor line exiting the exterior heat exchanger, which disadvantageously increases requisite defrost time, requiring more time for the interior air handler to remain totally un-functional, as Park does not simultaneously provide any interior heat (via simultaneously supplying hot gas to the interior air handler) while operating in his primary defrost cycle. Further, testing has indicated that, via switching from defrost to heating every 20-30 seconds, the heat supply ability of the interior air handler is materially impaired, since normal heat supply levels via the interior air handler are lost and take much longer than 20-30 seconds to build back up after the refrigerant flow to the interior air handler has been totally cut off during the preceding 20-30 second period.
Further, in Park's design, there is no disclosure that any extra refrigerant charge is required, which will impair optimum system operation since the addition of extra refrigerant transport lines, as taught by Park, will require some extra charge for optimum system performance; an optimum sizing of the hot gas line by-passing the indoor heat exchanger is not disclosed, which makes it impossible to know how to optimize defrost results; there is no mention of a timed operation of the fan in the exterior heat exchanger, and, of necessity as disclosed, will impose a power peaking period upon the system as a result of the defrost cycle operation as disclosed therein; and the periodic use of a liquid receiver (Park's 43) is required, which is an extra cost and expense that may be unnecessary. Also, it should be noted that Park's design does not call for a frequent cycle defrost mode of operation. Instead, it principally calls for an intense blast of 100% of the compressor's hot discharge gas, during the defrost cycle, traveling through the exterior heat exchanger (with no restriction in the exterior heat exchanger's exiting vapor line) during periods lasting from about 30-100 seconds, which periods are repeated multiple times at 20-30 second intervals when the frost is thick.